CN105814290B - Exhaust valve driving device and internal combustion engine having the exhaust valve driving device - Google Patents

Abstract

The invention provides an exhaust valve driving device which can change the closing time of an exhaust valve by a simple structure suitable for a marine internal combustion engine and adapt the characteristics of the internal combustion engine to the operating condition. An exhaust valve drive device (1) is provided with: a piston (7) provided on an exhaust valve (5) of an internal combustion engine; a cylinder (9) that houses the piston (7); a hydraulic pressure supply device (13) which is driven by a cam (35), intermittently supplies hydraulic pressure to the cylinder (9) at a predetermined valve opening timing, and presses the piston (7) to open the exhaust valve (5); an air spring device (15) that urges the exhaust valve (5) in the valve closing direction; a protrusion (7a) formed on the top surface of the piston (7) and having a top area smaller than the cross-sectional area of the piston (7); a recess (9a) formed on the inner top surface of the cylinder (9), the protrusion (7a) being inserted into the recess (9a) with a gap therebetween when the piston (7) is raised; an actuator (17) that changes the depth of the recess (9 a); and a control device (19) that controls the actuator (17).

Description

Exhaust valve driving device and internal combustion engine having the exhaust valve driving device

technical field

The present invention relates to a hydraulically actuated exhaust valve driving device which utilizes the oil of working oil ejected from a cam-driven plunger and an internal combustion engine provided with the exhaust valve driving device. The pressure pushes the piston provided at the end of the exhaust valve shaft to open the exhaust valve.

Background technique

This exhaust valve driving device can be controlled by the operating oil pressure, and the opening and closing of the exhaust valve can be kept optimal according to the operating load of the internal combustion engine.

For example, in a large marine low-speed two-stroke diesel engine, by delaying the closing time of the exhaust valve during high-load operation, the compression pressure of the gas in the cylinder can be prevented from being too high and the durability of the internal combustion engine can be improved. Furthermore, it is possible to reduce the closing speed of the exhaust valve and prevent the exhaust valve from colliding with the valve seat, thereby suppressing damage, wear, and the like of the exhaust valve and the valve seat.

However, in a general mechanical exhaust valve driving device in which the exhaust valve is directly driven by the cam, since the action of the exhaust valve depends on the profile of the cam, it is necessary to set a valve with a different profile when changing the opening and closing timing of the exhaust valve. A complex structure such as a plurality of cams or a rocker arm with a variable leverage ratio is provided between them. Therefore, it is not suitable for an exhaust valve driving device of a marine internal combustion engine which is intended to avoid failure at sea.

Patent Document 1 discloses a hydraulically actuated exhaust valve driving device in which the closing timing of the exhaust valve is delayed as described above. As shown in Fig. 1 of the same document, the piston 10 provided at the axial end of the exhaust valve 5 and pushing the exhaust valve 5 in the valve opening direction is a two-stage piston having a large diameter portion and a small diameter portion. The cylinder 4 that 10 slides is also a two-stage barrel shape with a large diameter and a small diameter.

The working oil pumped from the cam-driven hydraulic pump is supplied to the small bore of the cylinder 4 through the oil passage 11, and the piston 10 is pushed down by the oil pressure, and the exhaust valve 5 is opened. And, when the exhaust valve 5 is closed, the exhaust valve 5 is closed quickly until the small diameter portion of the piston 10 rushes into the small diameter of the cylinder 4, and when the small diameter portion of the piston 10 begins to rush into the small diameter of the cylinder 4, the sealing The working oil between the small-diameter part of the piston 10 and the large-diameter part of the cylinder 4 flows into the small-diameter part of the cylinder 4 through the gap between the cylinder 4 and the piston 10, and the flow resistance at this time has a buffering effect on the movement of the piston 10, The closing speed of the exhaust valve 5 decreases. Therefore, the exhaust valve 5 contacts the valve seat at a slower speed, and is free from the shock caused by the collision with the valve seat.

prior art literature

patent documents

Patent Document 1: Japanese Patent Laid-Open No. H1-244111

Contents of the invention

The problem to be solved by the invention

However, in the exhaust valve driving device described in Patent Document 1, the closing timing of the exhaust valve 5 is delayed to a certain extent, and for example, the closing timing of the exhaust valve 5 cannot be delayed or further delayed. Therefore, it is difficult to change the characteristics of the internal combustion engine to suit various operating conditions.

The present invention has been made in view of the above facts, and provides an exhaust valve driving device capable of changing the closing of the exhaust valve with a simple structure suitable for marine internal combustion engines, and an internal combustion engine provided with the exhaust valve driving device. time to adapt the characteristics of the internal combustion engine to the operating conditions.

Technical solutions

In order to solve the above-mentioned problems, the exhaust valve driving device and the internal combustion engine provided with the exhaust valve driving device according to the present invention employ the following methods.

A first aspect of the present invention is an exhaust valve driving device comprising: a piston provided on an exhaust valve of an internal combustion engine; a cylinder accommodating the piston; and a hydraulic supply mechanism communicating with the The cylinder is connected, driven by the cam and intermittently supplies oil pressure to the cylinder at the specified valve opening time, and pushes the piston to open the exhaust valve; the valve closing force mechanism, which controls the exhaust the valve is biased toward the valve closing direction; a convex portion is formed on the top surface of the piston and has a top area smaller than the cross-sectional area of the piston; a concave portion is formed on the inner top surface of the cylinder, When the piston rises, the convex portion is inserted into the concave portion with a gap; an actuator that changes the depth of the concave portion; and a control mechanism that controls the actuator, the cylinder has: a cylindrical first member; and a cylindrical second member disposed around the first member and relatively movable relative to the first member, the lower surface of the first member is located on the Above the lower surface of the second member, the recess is a space below the first member and on the inner peripheral side of the second member, and the actuator changes the recess by moving the second member depth.

Another aspect of the present invention is an exhaust valve driving device, which has: a piston installed on an exhaust valve of an internal combustion engine; a cylinder for accommodating the piston; It is connected with the cylinder, driven by the cam and intermittently supplies oil pressure to the cylinder at the specified valve opening time, and pushes the piston to open the exhaust valve; the air valve exerts force in a valve closing direction; a convex portion formed on the top surface of the piston and having a top area smaller than the cross-sectional area of the piston; a concave portion formed on the inner top surface of the cylinder, When the piston rises, the convex portion is inserted into the concave portion with a gap; an actuator that changes the depth of the concave portion; and a control mechanism that controls the actuator, the cylinder has: A cylindrical third member in the center, the lower surface of the third member is located below the inner top surface of the cylinder, and the recess is between the inner peripheral surface of the cylinder and the outer peripheral surface of the third member. space between, the actuator changes the depth of the recess by moving the third member.

According to the above configuration, when the hydraulic pressure supply mechanism is driven by the cam, hydraulic pressure is supplied to the cylinder at a predetermined valve opening timing, thereby pushing the piston inside the cylinder and opening the exhaust valve. And, when the oil pressure supplied to the cylinder decreases, the exhaust valve is closed by the biasing force of the valve closing biasing mechanism.

When the exhaust valve is closed, the piston that acts integrally with the exhaust valve enters the interior of the cylinder until the convex portion formed on the top surface of the piston is inserted into the concave portion formed on the internal top surface of the cylinder, and the exhaust valve moves faster Speed off. Then, when the convex part of the piston starts to rush into the concave part of the cylinder, the working oil sealed between the convex part and the cylinder flows out of the cylinder through the narrow gap between the convex part and the concave part. The action acts as a buffer, and the closing speed of the exhaust valve is reduced. Therefore, the exhaust valve contacts the valve seat at a slower speed and is free from the shock caused by the collision with the valve seat.

The depth of the concave portion of the cylinder can be changed by the actuator and the control mechanism. When the depth of the concave part is small, the amount of working oil sealed between the convex part and the cylinder is reduced, so the cushioning effect of the piston is reduced, and the exhaust valve will be at an earlier time close to the valve closing time specified by the profile of the cam. closure.

Also, when the depth of the concave portion is large, the amount of working oil sealed between the convex portion and the cylinder increases, so that it takes more time to discharge, and the cushioning effect of the piston increases. Consequently, the exhaust valve will close at a time later than the valve closing time dictated by the profile of the cam.

In this way, by changing the depth of the recess provided on the cylinder side, the closing timing of the exhaust valve can be advanced or delayed, so that the characteristics of the internal combustion engine can be adapted to the operating conditions.

Furthermore, since the change of the closing timing of the exhaust valve can be accomplished by a simple structure such as changing the depth of the concave portion provided on the cylinder, it is suitable for marine internal combustion engines that require a simple structure.

In the above configuration, the control means is preferably configured to control the actuator so that the depth of the recess increases as the load on the internal combustion engine increases.

According to this configuration, the closing timing of the exhaust valve is delayed as the load on the internal combustion engine increases. Accordingly, it is possible to prevent the compression pressure of the gas in the cylinder from being too high during high-load operation, thereby improving the durability of the internal combustion engine.

A second aspect of the present invention is an exhaust valve driving device comprising: a piston provided on an exhaust valve of an internal combustion engine; a cylinder accommodating the piston; The oil pressure is intermittently supplied to the cylinder when the valve is opened, and the piston is pushed to open the exhaust valve; the valve closing force mechanism applies force to the exhaust valve in the direction of closing the valve; the leakage path , which releases the oil pressure generated by the oil pressure supply mechanism; a flow regulating mechanism, which changes the passage area of the leakage passage; and a control mechanism, which controls the flow regulating mechanism.

According to the above configuration, as in the first aspect, when the oil pressure supply mechanism is driven by the cam, oil pressure is supplied to the cylinder at a predetermined valve opening timing, thereby pushing the piston inside the cylinder and opening the exhaust valve. And, when the oil pressure supply to the cylinder is interrupted, the exhaust valve is closed by the biasing force of the valve closing biasing mechanism.

A part of hydraulic oil (hydraulic pressure) supplied from the hydraulic supply mechanism leaks to the outside of the hydraulic circuit from the flow rate adjustment mechanism provided in the leakage passage. Therefore, when the exhaust valve is closed, the amount of oil flowing back from the cylinder to the hydraulic supply mechanism is less than the amount of oil that is sent to the cylinder when pressurized by the hydraulic supply mechanism. Therefore, when the exhaust valve is closed, the piston will It can be reliably returned to the cylinder, so that the exhaust valve can be reliably closed.

By controlling the flow rate adjustment mechanism, it is possible to adjust the amount of hydraulic fluid supplied from the hydraulic supply mechanism that leaks to the outside. When the leakage is small, the closing speed of the exhaust valve is slowed down, and when the leakage is large, the closing speed of the exhaust valve is accelerated.

In this way, by providing a flow regulating mechanism on the leakage passage that releases the hydraulic pressure generated by the hydraulic supply mechanism to the outside, the closing timing of the exhaust valve can be advanced or delayed, so that the characteristics of the internal combustion engine can be adapted to the operating conditions.

In addition, since the change of the closing timing of the exhaust valve can be accomplished with a simple structure such as providing a leakage passage and a flow rate adjustment mechanism on a hydraulic circuit or a cylinder, it is suitable for marine internal combustion engines that require a simple structure.

In the above configuration, the control means is preferably configured to control the flow rate adjustment means so that the passage area of the leakage passage decreases as the load on the internal combustion mechanism increases.

According to this configuration, the closing timing of the exhaust valve is delayed as the load of the internal combustion engine increases, and the durability of the internal combustion engine can be improved by preventing the compression pressure of the cylinder gas from being too high during high-load operation.

A third aspect of the present invention is an internal combustion engine including the exhaust valve driving device according to any one of the above.

Thus, a simple structure suitable for marine internal combustion engines can be formed, and the closing timing of the exhaust valve can be changed, so that the characteristics of the internal combustion engine can be adapted to the operating conditions.

Beneficial effect

As described above, according to the exhaust valve drive device and the internal combustion engine having the exhaust valve drive device according to the present invention, the closing timing of the exhaust valve can be changed with a simple structure suitable for a marine internal combustion engine, whereby the characteristics of the internal combustion engine can be changed. Adapt to the operating conditions, and can improve the reliability and durability of the internal combustion engine, while also helping to save fuel, etc.

Description of drawings

FIG. 1 is a schematic configuration diagram showing an exhaust valve driving device according to a first embodiment of the present invention.

Fig. 2 is an enlarged view of part II in Fig. 1, (a) is a longitudinal sectional view showing the state before the convex part of the piston is inserted into the concave part of the cylinder, and (b) is a state showing the state where the convex part of the piston starts to be inserted into the concave part of the cylinder Longitudinal section view.

Fig. 3 is a diagram showing other shape examples of the convex portion of the piston and the concave portion of the cylinder, (a) is a longitudinal sectional view showing the state before the convex portion of the piston is inserted into the concave portion of the cylinder, and (b) is a diagram showing the start of the convex portion of the piston. A longitudinal sectional view of a state where the cylinder is inserted into the recess.

4( a ), ( b ), and ( c ) are graphs respectively showing the cam lifting amount, the operating hydraulic pressure, and the exhaust valve lifting amount in the first embodiment.

5 is a schematic configuration diagram showing an exhaust valve driving device according to a second embodiment of the present invention.

6( a ), ( b ), and ( c ) are graphs respectively showing the cam lifting amount, the operating hydraulic pressure, and the exhaust valve lifting amount in the second embodiment.

Detailed ways

Embodiments of the exhaust valve driving device according to the present invention will be described below with reference to the drawings.

[the first embodiment]

FIG. 1 is a schematic configuration diagram showing an exhaust valve driving device according to a first embodiment of the present invention. This exhaust valve driving device 1 is installed in a diesel engine (internal combustion engine) for a ship main engine.

Diesel engines for marine main engines (hereinafter referred to as "diesel engines") are, for example, low-speed two-stroke engines, and employ a unidirectional scavenging single-flow type in which air is supplied from below to exhaust from above. The output of the diesel engine is directly or indirectly connected to the propeller via a propeller shaft not shown in the figure.

As shown in FIG. 1 , the exhaust valve driving device 1 has: an exhaust valve 5 that switches the exhaust flow path formed on the cylinder head 3 of the diesel engine; a piston 7 that is provided on the exhaust valve 5; 9, which accommodates the piston 7; the hydraulic circuit 11 and the hydraulic supply device (oil pressure supply mechanism) 13, which supplies oil pressure to the cylinder 9; the air spring device (valve closing force mechanism) 15, which controls the exhaust valve 5 Apply force to the valve closing direction (upper in FIG. 1 ); the actuator 17 and the control device (control mechanism) 19 .

The piston 7 is connected to the upper end of the shaft portion 5 a of the exhaust valve 5 extending in the vertical direction, and reciprocates in the vertical direction within the cylinder 9 as the exhaust valve 5 is opened and closed. A hydraulic chamber 21 formed by the piston 7 and the cylinder 9 is connected to one end 11 a of the hydraulic circuit 11 . Further, a throttling circuit 25 extends from the oil pressure chamber 21 , and a throttling device 27 as a fixed throttle is provided on the throttling circuit 25 . In addition, the exhaust valve 5 is always biased in the valve closing direction (upward) by the air spring device 15 .

The hydraulic supply device 13 has a structure including a plunger 31 , a cylinder 33 and a cam 35 . The punch rod 31 can be freely slidably inserted into the cylinder 33, and the pressurized chamber 37 is connected with the other end 11b of the hydraulic circuit 11. The pressurized chamber 37 is formed by the punch rod 31 and the cylinder 33, and the punch rod 31 is always pressed A urging mechanism not shown in the figure urges in a direction (downward) away from the air cylinder 33 .

At the lower portion of the punch 31 , a cam roller 41 is supported via a connecting shaft 39 . The cam roller 41 rotates along the outer peripheral surface of the cam 35 arranged below, that is, the cam profile. The cam 35 is provided integrally with a camshaft 43 which rotates synchronously with the crankshaft of the diesel engine.

In the hydraulic circuit 11, a low-pressure hydraulic oil supply circuit 45 is branched from a branch point 11c. A low-pressure hydraulic oil source (not shown) is connected to the low-pressure hydraulic oil supply circuit 45 via a check valve 47 , and oil pressure used as a reference when opening and closing the exhaust valve 5 is supplied. When the hydraulic pressure in the hydraulic circuit 11 is below a predetermined value, the check valve 47 is opened, and hydraulic oil (oil pressure) is replenished from the low-pressure hydraulic oil supply circuit 45 . And, when the pressure in the hydraulic circuit 11 reaches a predetermined value or more, that is, when the pressurization process is performed by the plunger 31, the check valve 47 is closed.

As shown in FIGS. 2( a ) and ( b ), a cylindrical convex portion 7 a is formed at the center of the top surface of the piston 7 . The top area of the convex portion 7 a is smaller than the cross-sectional area of the piston 7 . For example, assuming that the diameter of the piston 7 is 80 mm, the diameter of the convex portion 7a shown in the example may be about 50 mm, and the height of the convex portion 7a from the top surface is about 60 mm, but it is not limited to this size or this diameter ratio.

In addition, a cylindrical hole-shaped recess 9 a is formed in the center of the inner top surface of the cylinder 9 . The inner diameter of the concave portion 9a is set to a size that the convex portion 7a of the piston 7 is inserted with a gap of about several millimeters when the piston 7 is raised.

The depth h of the recess 9a can range from zero to the same height as the protrusion 7a. For example, the structure of the top of the air cylinder 9 is a structure in which the cylindrical movable member 9c is tightly arranged around the cylindrical fixed member 9b positioned at the central part and can move up and down relative to each other. The space on the inner peripheral side of the member 9c constitutes the recess 9a. Moreover, the hydraulic circuit 11 opens to the lower surface of the fixing member 9b.

Furthermore, by moving the movable member 9c up and down by the actuator 17 shown in FIG. 1, the depth h of the recessed part 9a can be changed. For example, it is conceivable to make a screw pair between the outer peripheral surface of the fixed member 9b and the inner peripheral surface of the movable member 9c, and use the power of the actuator 17 to make the movable member 9c rotate relative to the fixed member 9b, thereby The movable member 9c is moved up and down to change the depth of the recessed portion 9a. In addition, the space between the inner peripheral surface of the air cylinder 9 and the outer peripheral surface of the movable member 9 c may be provided as a screw pair.

Furthermore, the control device 19 shown in FIG. 1 controls the actuator 17 to set the vertical position of the movable member 9c. For example, the control device 19 controls the actuator 17 so that the depth of the recessed portion 9a increases as the load on the diesel engine increases.

Next, the operation of the exhaust valve driving device 1 configured as described above will be described.

When the cam 35 (cam shaft 43) of the oil pressure supply device 13 rotates, the cam roller 41 moves up and down while rotating along the cam profile of the cam 35, and the up and down movement causes the punch 31 to move up and down in the cylinder 33 through the connecting shaft 39. slide.

When the punch 31 slides upward in the cylinder 33 , the working oil filled in the pressurized chamber 37 is pressurized, and the working oil is sent to the oil between the cylinder 9 and the piston 7 through the hydraulic circuit 11 . In the pressure chamber 21. The volume of the hydraulic chamber 21 is enlarged by the hydraulic pressure of the working oil, and the piston 7 is pushed down against the urging force of the air spring device 15 to open the exhaust valve 5 . The opening amount of the exhaust valve 5 is determined by the height of the cam 35 from the reference circle 35a.

And, when the cam 35 rotates downward, the plunger 31 is pressed back downward by the force applying mechanism not shown in the figure, and the oil pressure applied to the pressurizing chamber 37 and the oil pressure chamber 21 drops to the level supplied from the low-pressure working oil. The weaker reference oil pressure supplied by circuit 45. Therefore, the exhaust valve 5 is pushed and closed by the biasing force of the air spring device 15, whereby the piston 7 rises, the volume of the hydraulic chamber 21 becomes minimum, and the working oil in the hydraulic chamber 21 flows back through the hydraulic circuit 11. To the pressurized chamber 37 of the cylinder 33 .

In this way, the oil pressure supply device 13 is driven by the cam 35 to intermittently supply oil pressure to the cylinder 9 at a predetermined valve opening timing, thereby pushing the piston 7 to open the exhaust valve 5 .

Furthermore, when pressurization is performed by the plunger 31 , a small amount of working oil in the hydraulic chamber 21 is discharged from the throttling device 27 of the throttling circuit 25 to the outside of the hydraulic circuit 11 . Therefore, when the exhaust valve 5 is closed, the amount of oil flowing back from the oil pressure chamber 21 to the pressurization chamber 37 is less than the amount of oil delivered from the pressurization chamber 37 to the oil pressure chamber 21 when pressurized by the plunger 31. The amount can make the piston 7 rise to the uppermost part of the cylinder 9 and close the exhaust valve 5 reliably. When the plunger 31 is not pressed by the cam 35 , the amount of working oil discharged from the throttle device 27 is supplied from the low-pressure working oil supply circuit 45 to the hydraulic circuit 11 .

FIG. 4 is a graph showing the relationship between the lift amount (a) of the cam 35 , the operating hydraulic pressure (b) in the hydraulic chamber 21 , and the lift amount (c) of the exhaust valve 5 . In FIGS. 4( b ) and ( c ), the lines indicated by solid lines represent the operating oil pressure and the exhaust valve lifting amount when the depth h of the recessed portion 9 a shown in FIG. 2( a ) is zero.

At time t0, when the cam lift increases with the profile of the cam 35 and the plunger 31 starts to lift, the working oil pressure of the oil pressure chamber 21 starts to rise from the reference pressure. At time t1, the amount of cam lifting reaches the maximum value, the plunger 31 is pushed up to the top dead center, and the actuating oil pressure reaches the maximum value. At this time, at time t2, the oil pressure in the oil pressure chamber 21 overcomes the applied force of the air spring device 15 And cylinder internal pressure and piston 7 is depressed.

As a result, the lifting amount of the exhaust valve 5 is increased, and the exhaust valve 5 is fully opened at time t3. At this time, the volume of the oil pressure chamber 21 increases as the piston 7 is depressed. Therefore, although the oil pressure in the oil pressure chamber 21 decreases sharply, the oil pressure required to open the exhaust valve 5 can be maintained. Therefore, while the plunger 31 is maintained at the top dead center following the profile of the cam 35, the lifting amount of the exhaust valve 5 is also maintained at the maximum, and the exhaust valve 5 remains in the open state.

At time t5, when the cam lifting amount decreases with the profile of the cam 35 and the plunger 31 starts to descend, the operating oil pressure of the oil pressure chamber 21 also begins to descend. When the operating oil pressure drops to a predetermined value, the applied force of the air spring device 15 and the pressure in the cylinder prevail, and the piston 7 is pushed upward from time t6, so that the lifting amount of the exhaust valve 5 begins to decrease. When the lifting amount of the cam 35 is zero and the plunger 31 is lowered to the bottom dead center, the exhaust valve 5 is completely closed at time t7. Then, the working hydraulic pressure of the hydraulic circuit 11 returns to the reference pressure.

As shown in Figure 2(a), when the movable member 9c descends relative to the fixed member 9b of the cylinder 9 to form a recess 9a of depth h, if the exhaust valve 5 is closed, the piston 7 that operates integrally with the exhaust valve 5 Entering the inside of the cylinder 9 , the volume of the hydraulic chamber 21 decreases, and the working oil filled in the hydraulic chamber 21 is released from the hydraulic circuit 11 and flows back into the pressurizing chamber 37 .

At this time, since the working oil in the hydraulic chamber 21 will smoothly flow into the hydraulic circuit 11 until the convex portion 7a of the piston 7 is inserted into the concave portion 9a of the cylinder 9, the piston 7 enters the interior of the cylinder 9 at a relatively fast speed, and is discharged. Air valve 5 is closed quickly. Then, as shown in FIG. 2( b ), when the convex portion 7a starts to be inserted into the concave portion 9a, the hydraulic chamber 21 is divided into a chamber 21a formed around the convex portion 7a and a chamber 21b formed inside the concave portion 9a.

The operating oil in the chamber 21b is smoothly discharged from the hydraulic circuit 11 as it is, while the operating oil sealed in the chamber 21a flows into the chamber 21b through the narrow gap between the chamber 21a and the chamber 21b, and then passes through the hydraulic circuit 11. discharge. Therefore, as the working oil passes through the gap, the huge flow resistance has a cushioning effect on the action of the piston 7 (Cushioning Action), which reduces the valve closing speed of the exhaust valve 5 and delays until the exhaust valve 5 is completely closed. moment.

If the depth h of the recessed portion 9a is small, the required amount of oil flowing from the chamber 21a into the chamber 21b is reduced, and thus the time for applying a buffering action to the upward movement of the piston 7 is shortened. Consequently, the exhaust valve 5 will close at an earlier moment closer to the valve closing moment dictated by the profile of the cam 35 .

Also, as the depth h of the recessed portion 9a increases, since the required amount of oil flowing from the chamber 21a into the chamber 21b increases, it takes more time to discharge, and the time for buffering the upward movement of the piston 7 increases. Therefore, the exhaust valve 5 closes at a timing greatly delayed from the valve closing timing specified by the profile of the cam 35 .

In this way, when the rising process of the piston 7 is about to be terminated and the convex portion 7a of the piston 7 is inserted into the concave portion 9a of the cylinder 9, the hydraulic chamber 21 (chamber 21a) will generate flow resistance due to the working oil sealed in the chamber 21b. ) The pressure in Figure 4(b) rises sharply as shown by the dotted lines P1 and P2. A dotted line P1 indicates a pressure increase rate when the depth h of the recessed portion 9a is small, and a dotted line P2 indicates a pressure increase rate when the depth h of the recessed portion 9a is large.

In this way, since the pressures P1 and P2 in the oil pressure chamber 21 rise sharply, as shown by the dotted lines L1 and L2 in Fig. 4(c), when the exhaust valve 5 is about to close, the reduction rate of the lifting amount A slow tilt occurs. The reduction rate of the lifting amount of the exhaust valve 5 is a portion indicated by a dotted line L1 when the pressure in the hydraulic chamber 21 is P1, and a portion indicated by a dotted line L2 when the pressure in the hydraulic chamber 21 is P2. That is, the greater the depth h of the recessed portion 9a, the longer the time required until the exhaust valve 5 is completely closed (the valve closing timing is delayed). Therefore, the exhaust valve 5 will contact the valve seat (valve seat) at a slower speed, and is free from the shock caused by the collision with the valve seat.

In this way, by changing the depth of the recessed portion 9a provided on the side of the cylinder 9, the closing timing of the exhaust valve 5 can be made close to or later than the valve closing timing specified by the cam profile of the cam 35, so that the characteristics of the diesel engine can be adapted. operating condition.

And, the change of the closing timing of the exhaust valve 5 can be done by changing the depth of the recess 9a provided on the cylinder 9, that is, by using the actuator 17 to make the movable member 9c relatively move in the axial direction with respect to the fixed member 9b of the cylinder 9. Since it can be completed with a simple structure, it can be used as a structure suitable for a marine diesel engine that wants to use a simple structure.

In addition, the control device 19 controls the actuator 17 so that the depth h of the recessed portion 9a increases as the load on the diesel engine increases. Therefore, as the load on the diesel engine increases, the closing timing of the exhaust valve 5 is delayed. Accordingly, it is possible to prevent the compression pressure of the gas in the cylinder from being too high during high-load operation, thereby improving the durability of the diesel engine.

3( a ), ( b ) are longitudinal cross-sectional views showing other shape examples of the convex portion 7 a of the piston 7 and the concave portion 9 a of the cylinder 9 .

Here, the convex portion 7 a provided on the top surface of the piston 7 is formed to protrude in a cylindrical shape from the periphery of the top surface of the piston 7 . In addition, the recessed portion 9a provided on the inner top surface of the cylinder 9 is formed as a cylindrical recess around the inner top surface. That is, the radially inner and outer positional relationship between the convex portion 7 a and the concave portion 9 a shown in FIGS. 2( a ) and ( b ) is reversed.

The depth h of the recess 9a can range from zero to the same height as the protrusion 7a. For example, a columnar movable member 9d arranged at the top center of the cylinder 9 can move up and down. When the movable member 9d protrudes from the inner top surface of the cylinder 9, there Recesses 9a are formed between the outer peripheral surfaces.

The outer diameter of the movable member 9d is set such that when the piston 7 rises, the protrusion 7a of the piston 7 surrounds the periphery of the movable member 9d with a gap of about several millimeters. The movable member 9d is driven up and down by the actuator 17 shown in FIG. 1, thereby changing the depth h of the recess 9a. Furthermore, the hydraulic circuit 11 opens to the lower surface of the movable member 9d.

As shown in Figure 3 (a), when the movable member 9d of the cylinder 9 descends to form a recess 9a of depth h, if the exhaust valve 5 is closed, the piston 7 that operates integrally with the exhaust valve 5 enters into the cavity of the cylinder 9. Internally, the volume of the hydraulic chamber 21 is thereby reduced, and the working oil filled in the hydraulic chamber 21 is released from the hydraulic circuit 11 and flows back into the pressurizing chamber 37 .

At this time, since the working oil in the hydraulic chamber 21 will smoothly flow into the hydraulic circuit 11 until the convex portion 7a of the piston 7 is inserted into the concave portion 9a of the cylinder 9, the piston 7 enters the interior of the cylinder 9 at a relatively fast speed, and is discharged. Air valve 5 is closed quickly. Then, as shown in FIG. 3( b), when the convex portion 7a starts to be inserted into the concave portion 9a, the oil pressure chamber 21 is divided into a chamber 21a formed on the inner peripheral side of the convex portion 7a and a chamber formed inside the concave portion 9a. 21b.

The operating oil in the chamber 21a is smoothly discharged from the hydraulic circuit 11 as it is, while the operating oil sealed in the chamber 21b flows into the chamber 21a through the narrow gap between the chamber 21b and the chamber 21a, and then passes through the hydraulic circuit 11. discharge. Therefore, as the working oil passes through the gap, the huge flow resistance has a cushioning effect on the action of the piston 7 (Cushioning Action), which reduces the valve closing speed of the exhaust valve 5 and delays until the exhaust valve 5 is completely closed. moment. In this way, the exhaust valve 5 can be protected from the impact caused by the collision with the valve seat, and the characteristics of the diesel engine can be adapted to the operating conditions.

As in the case of the structure in Fig. 2(a) and (b), the greater the depth h of the recessed portion 9a, the longer the time for the cushioning action (Cushioning Action) by the flow resistance of the working oil to act. That is, the valve closing timing of the exhaust valve 5 can be delayed as the depth h of the recessed portion 9 a is larger.

According to the structure of this FIG. 3 (a), (b), the structure which changes the height of the recessed part 9a can be made simpler than the structure shown in FIG. 2 (a), (b).

[the second embodiment]

5 is a schematic configuration diagram showing an exhaust valve driving device according to a second embodiment of the present invention. In this exhaust valve driving device 51, the difference from the exhaust valve driving device 1 of the first embodiment is that there are no convex parts on the top surface of the piston 7 and concave parts on the inner top surface of the cylinder 9; Leak passage 53 is branched at branch point 11d of circuit 11, and variable throttle device (flow rate adjustment mechanism) 55 capable of changing the passage area of leakage passage 53 is connected therebetween. Since the configuration and function of other parts are the same as those of the exhaust valve driving device 1 of the first embodiment, the same reference numerals are assigned to the respective parts and corresponding descriptions are omitted.

The throttling amount of the variable throttle device 55 is controlled by a control device (control means) 57 . The control device 57 controls the throttling amount of the variable throttle device 55 so that the passage area of the leakage passage 53 decreases as the load on the diesel engine increases. In addition, instead of the variable throttle device 55, a flow rate regulating valve or the like may be provided.

According to this exhaust valve driving device 51, similar to the exhaust valve driving device 1 of the first embodiment, when the hydraulic pressure supply device 13 is driven by the cam 35, hydraulic pressure is supplied to the cylinder 9 at a predetermined valve opening timing, thereby The piston 7 inside the cylinder 9 is pushed and the exhaust valve 5 is opened. And, when the oil pressure supply to the cylinder 9 is interrupted, the exhaust valve 5 is closed by the urging force of the air spring device 15 .

Part of the hydraulic oil (hydraulic pressure) supplied from the hydraulic supply device 13 leaks to the outside of the hydraulic circuit 11 through the variable throttle device 55 provided in the leakage passage 53 . Therefore, the amount of oil flowing back from the cylinder 9 to the hydraulic supply device 13 when the exhaust valve 5 is closed is smaller than the amount of oil that is pressure-fed to the cylinder 9 when the hydraulic supply device 13 pressurizes it. Therefore, when the exhaust valve 5 is closed, the piston 7 can be raised to the uppermost position in the cylinder 9 and the exhaust valve 5 can be closed reliably.

By controlling the variable throttle device 55 , it is possible to adjust the amount of hydraulic fluid supplied from the hydraulic supply device 13 that leaks to the outside. When the leakage amount is small, the closing speed of the exhaust valve 5 is slowed down, and when the leakage amount is large, the closing speed of the exhaust valve 5 is accelerated.

In this way, the closing timing of the exhaust valve 5 can be advanced or delayed by providing the variable throttling device 55 on the leakage passage 53, which is supplied from oil pressure. The hydraulic circuit 11 between the device 13 and the cylinder 9 is branched out.

Furthermore, since the change of the closing timing of the exhaust valve 5 can be accomplished with a simple structure such as providing the leakage passage 53 and the variable throttle device 55 in the hydraulic circuit 11, it is suitable for marine diesel engines that require a simple structure. In addition, the leakage passage 53 does not necessarily have to branch from the hydraulic circuit 11 , and may branch from the cylinder 9 , for example.

6 is a graph showing the relationship between the lift amount (a) of the cam 35 on the exhaust valve driving device 51 , the hydraulic pressure in the hydraulic chamber 21 (b) and the lift amount (c) of the exhaust valve 5 . Since the basic operation is the same as that of the exhaust valve driving device 1 according to the first embodiment described with reference to FIG. 4 , repeated description will be omitted.

When the throttling amount of the variable throttling device 55 of the leakage passage 53 is the minimum and the exhaust valve 5 is closed, since the operating oil in the hydraulic chamber 21 flows back to the hydraulic supply device 13 and does not leak much, Therefore it will be more time consuming for the piston 7 to rise in the cylinder 9 . Therefore, as shown by the solid line in FIG. 6( b ), the reduction of the operating hydraulic pressure in the hydraulic chamber 21 corresponds to the reduction in the lift amount of the cam 35 . Furthermore, the decrease in the lift amount of the exhaust valve 5 is shown by the solid line in FIG. 6( c ), and the closing timing of the exhaust valve 5 is delayed.

In addition, when the throttling amount of the variable throttle device 55 is increased, the leakage amount of the working oil in the hydraulic chamber 21 increases, so the amount of working oil flowing back from the hydraulic chamber 21 to the hydraulic supply device 13 will be reduced, and the rising time of the piston 7 in the cylinder 9 will be shortened. Therefore, as shown by the dotted line in FIG. 6( b ), the operating oil pressure in the oil pressure chamber 21 decreases faster than the decrease in the lift amount of the cam 35 . Furthermore, the decrease in the lift amount of the exhaust valve 5 is shown by the dotted line in FIG. 6( c ), and the closing timing of the exhaust valve 5 is advanced.

As described above, the control device 57 controls the throttling amount of the variable throttle device 55 so that the passage area of the leakage passage 53 decreases as the load on the diesel engine increases. Therefore, as the load on the diesel engine rises, the closing timing of the exhaust valve 5 is delayed, so that the compression pressure of the cylinder gas during high-load operation can be prevented from being too high and the durability of the diesel engine can be improved.

In addition, as an improvement, as shown by the dotted line in FIG. 5 , a leakage passage 53a connected between the hydraulic chamber 21 and the hydraulic circuit 11 may be provided, and a variable throttle device 55a may be provided on the leakage passage 53a, so that The rising time of the piston 7 when the exhaust valve 5 is closed can be adjusted.

As described above, according to the exhaust valve driving device 1, 51 and the diesel engine provided with the exhaust valve driving device according to the embodiment of the present invention, it is possible to change the position of the exhaust valve 5 with a simple structure suitable for a marine diesel engine. The closing time adapts the characteristics of the diesel engine to the operating conditions, and can improve the reliability and durability of the internal combustion engine, while also helping to save fuel, etc.

In addition, the present invention is not limited only to the configuration of the above-mentioned embodiments, but can be appropriately changed or improved within the range not departing from the gist of the present invention, and the embodiments after such changes or improvements are also included in the scope of rights of the present invention. .

Symbol Description

1.51 Exhaust valve driving device

5 exhaust valve

7 pistons

7a Convex

9 cylinders

9a Recess

11 Hydraulic circuit

13 Hydraulic supply device (hydraulic supply mechanism)

15 Air spring device (valve closing force mechanism)

17 Actuators

19, 57 control device (control mechanism), and

21 Hydraulic chamber

31 Punch

35 cam

37 pressurized chamber

53 leak path

55 Variable throttling device (flow adjustment mechanism)

Claims (6)

1. a kind of exhaust valve actuator, which is characterized in that have：Piston is arranged on the air bleeding valve of internal combustion engine；Cylinder, It accommodates the piston；Oil pressure feed mechanism is connect via oil hydraulic circuit with the cylinder, by actuated by cams and defined The valve opening moment, and pushing the piston made the air bleeding valve open intermittently to the cylinder for oil feed pressure；Valve closing force machine Structure exerts a force to the air bleeding valve to valve closing direction；Protrusion is formed on the top surface of the piston and with than the piston The small topside area of cross-sectional area；Recess portion is formed in the inside top surface of the cylinder, when the piston rises, institute It states protrusion and is separated with and be inserted into gap in the recess portion；Actuator changes the depth of the recess portion；And control mechanism, The actuator is controlled,

The cylinder has：Positioned at the columned first component of central part；And be arranged around the first component and Relative to the second component of the relatively-movable cylindrical shape of the first component,

The lower surface of the first component is located at the top of the lower surface of the second component, and the recess portion is the first component Lower section, the space of the inner circumferential side of the second component,

The actuator changes the depth of the recess portion by the movement second component.

2. a kind of exhaust valve actuator, which is characterized in that have：Piston is arranged on the air bleeding valve of internal combustion engine；Cylinder, It accommodates the piston；Oil pressure feed mechanism is connect via oil hydraulic circuit with the cylinder, by actuated by cams and defined The valve opening moment, and pushing the piston made the air bleeding valve open intermittently to the cylinder for oil feed pressure；Valve closing force machine Structure exerts a force to the air bleeding valve to valve closing direction；Protrusion is formed on the top surface of the piston and with than the piston The small topside area of cross-sectional area；Recess portion is formed in the inside top surface of the cylinder, when the piston rises, institute It states protrusion and is separated with and be inserted into gap in the recess portion；Actuator changes the depth of the recess portion；And control mechanism, The actuator is controlled,

The cylinder has：Positioned at the columned third component of central part,

The lower surface of the third component is located at the lower section of the inside top surface of the cylinder, and the recess portion is the inner circumferential of the cylinder Space between face and the peripheral surface of the third component,

The actuator changes the depth of the recess portion by the movement third component.

3. exhaust valve actuator according to claim 1 or 2, which is characterized in that held described in the control mechanism control Row device makes the depth of the recess portion increase with the rising of the load of the internal combustion engine.

4. a kind of exhaust valve actuator, which is characterized in that have：Piston is arranged on the air bleeding valve of internal combustion engine；Cylinder, It accommodates the piston；Oil pressure feed mechanism is connect via oil hydraulic circuit with the cylinder, by actuated by cams and defined The valve opening moment, and pushing the piston made the air bleeding valve open intermittently to the cylinder for oil feed pressure；Valve closing force machine Structure exerts a force to the air bleeding valve to valve closing direction；Leakage path, from the connection oil pressure feed mechanism and the cylinder Branch in the oil hydraulic circuit and go out；Flow control device changes the area of passage of the leakage path；And control machine Structure controls the flow control device, wherein the control mechanism controls the flow control device, makes the air bleeding valve The oil mass backflowed from the cylinder to the oil pressure feed mechanism when valve closing is less than through the oil pressure feed mechanism for oil feed pressure When supply to the oil mass of the cylinder.

5. exhaust valve actuator according to claim 4, which is characterized in that the control mechanism controls the flow tune Mechanism is saved, the area of passage of the leakage path is made to reduce with the rising of the load of the internal combustion engine.

6. a kind of internal combustion engine, which is characterized in that with the exhaust valve actuator described in any one of claim 1,2,4 or 5.